1. Field of the Invention
The present invention generally relates to heat exchangers and particularly to condensers for maintaining high heat transfer during a phase change of a fluid being cooled under adverse inertial loading.
2. Description of the Prior Art
Devices for providing a transfer of heat to or from a fluid have long been known. Similarly, devices for transferring heat between two fluids while largely preventing mixing or even contact between the fluids are familiar to most people. An automobile radiator would be a commonly recognized examples of such a device.
More modern developments in environmental controls, refrigeration and other thermal control systems and energy management systems have prompted the development of arrangements capable of absorbing heat at one location and discharging it at another through the use of a fluid commonly referred to as a refrigerant which may or may not undergo a phase change in the process. As is well known, the basic elements of such a system include a compressor which raises the pressure and, hence, the temperature of the refrigerant, a condenser which removes the heat from the compressed refrigerant and typically causes a phase change of a portion of the mass of the refrigerant from a gas to a liquid, and an evaporator in which heat is absorbed during reversion of the refrigerant from the liquid phase to the vapor phase.
Various arrangements which effectively increase the surface area of a heat exchanger will improve the thermal performance thereof. Direction of a fluid over the surface of such a device will also enhance the possible rate of heat exchange for a given temperature differential between the heat exchanger and the fluid. It has been recently found that the direction of a jet of fluid at a surface of the heat exchanger provides a beneficial degree of turbulence to enhance heat transfer. This has led to the development of the jet fin type of heat exchanger where a plurality of apertured plates are provided in a generally parallel arrangement. The aperture locations in sequential ones of the plates are skewed from each other so that each aperture forms a jet which directs a fluid against a solid portion of the next plate. U.S. Pat. Nos. 4,494,171, to Bland et al. and 4,880,055, to Niggemann et al, both assigned to the assignee of the present invention, are typical of jet fin heat exchanger structures.
While jet fin heat exchangers exhibit enhanced heat transfer coefficients, and enhanced heat transfer surface area per unit volume of the heat exchanger, they are not well suited to a condenser structure since the existence of plural phases of a fluid will interfere with flow within the jets due to differences in viscosity of the liquid and vapor phases. While the geometry of a jet fin heat exchanger can be reasonably optimized for either a liquid or a vapor phase fluid (e.g. a small range of viscosity) to be circulated therethrough, two phases of greatly differing viscosity cannot be efficiently accommodated and, in practice, jet fin heat exchangers are typically designed only for liquid phase fluids in aircraft applications.
In the condenser element, with which the present invention is principally concerned, heat transfer from the high-temperature compressed gas to the condenser structure an thence to another fluid may be impeded by the phase and the nature of the refrigerant flow within refrigerant-containing passages of the condenser. Generally, the heat transfer rate is limited by the thickness of condensed liquid phase refrigerant on the passage walls which will change with the type of flow (e.g. flow regime) which is present in the passages. The relative amount of vapor in the total fluid mass (vapor and liquid) is generally referred to as the quality of the fluid in a given volume thereof and will affect the heat transfer rate, as well. The presence of non-condensible gases will also reduce the efficiency of the condenser by displacing condensible vapor. The heat transfer rate will also be affected by the rate of liquid phase flow which must, in turn, be propelled by the vapor phase flow or by gravity. Types of two-phase flow generally encountered in condensers, in order of generally decreasing heat transfer coefficient are as follows:
Mist annular flow--same as annular flow with condensed liquid phase refrigerant droplets which are pulled off the liquid phase surface and entrained in the vapor phase, reducing the thickness of the liquid annulus; PA1 Annular flow--flow during which all interior surfaces of the passages are wetted with condensed vapor, thus presenting an annular cross-section of liquid phase refrigerant; PA1 Semi-annular flow--similar to annular flow but where the thickness of the liquid phase layer is not uniform and possibly great enough for the existence of wave-like turbulence to exist due to the vapor velocity; PA1 Stratified flow--a typically gravity dominated flow where not all interior surfaces of the passages are wetted by the liquid phase refrigerant (e.g. liquid phase refrigerant collects at only a portion of the interior passage surface circumference)due to the forces of the gravity field; and PA1 Slug flow--flow where liquid phase of the refrigerant collects at only a portion of the passage length, leaving some interior passage surface areas relatively unwetted with liquid and vapor phases alternately filling the passage cross-section. Slug flow in a normal gravity field would appear as a wavy stratified flow with a majority of wave tops reaching the tops of the conduit to fill the cross-section thereof. In a low gravity field, shear forces will dominate and the flow will resemble annular flow in some portions of the conduit and plug flow in others. PA1 Plug Flow--similar to slug flow at particularly low values of "quality"; the ratio of vapor phase mass to liquid phase mass. In a normal gravity field, this would appear as a largely fluid filled conduit with bubbles of vapor phase near an upper boundary of the conduit. In a low gravity field the bubbles would be entrained in the liquid, but of such small size that heat transfer would resemble that of a liquid-filled conduit due to the thickness of the liquid phase layer surrounding the bubbles.
Slug flow is generally classified into two types. So-called high velocity slug flow is generally shear force dominated and so-called low velocity slug flow is generally gravity field dominated. This terminology is used to indicate the relative dominance of shear forces or gravity field forces since, in a normal gravity field (e.g. approximately 1 g), the relative dominance of these forces will depend upon the velocity of the flow. However, in a reduced gravity field, as would be encountered in spacecraft, a much reduced velocity would result in so-called high velocity type slug flow. At high flow velocities or low gravity field, bullet-shaped Taylor bubbles will form due to the dominance of shear forces. As indicated above, the flow regime at the location of a Taylor bubble will resemble annular flow and locations between bubbles will resemble liquid flow. Heat transfer in a so-called high velocity slug flow can be approximated by pro-rating the heat transfer of these two flow regimes over the length of the conduit based on the proportion of the length of the conduit containing Taylor bubbles to the length filled with liquid phase fluid. Generally, heat transfer will remain good in a high velocity slug flow regime.
As can be readily appreciated, the above order corresponds to increasing thickness, decreasing area of the liquid phase wetting the interior surface of the passage and/or reduced ability of the liquid phase to be driven through the passage by the vapor phase. All of these conditions adversely affect the efficiency of heat transfer from the two-phase fluid to the walls of the passage. Therefore, it is seen that the type of flow within a condenser passage containing a refrigerant in two phases can severely affect the performance of the condenser and the system in which it is installed. By the same token, the above order of flow regimes reflects decreasing shear force dominance and increasing gravity or inertial force dominance. Accordingly, alteration of the flow regime due to changes in acceleration of a condenser can have serious adverse effects on the performance of the condenser, as will now be explained.
Referring briefly now to FIG. 7, a flow regime map assuming a normal gravity field of 1 g is illustrated. The flow pattern or regime in a channel or conduit depends on several parameters including flow rate, quality, fluid properties and heat transfer. In order to determine analytically what flow regime the condenser is in, a two-phase flow regime parameter is calculated. Also, the quality of the fluid will change over the length of a condenser passage and the shear forces within the fluid will be different for the liquid and vapor phases, resulting in different pressure drops within the conduit over an incremental portion of the length of the conduit. It is therefore useful to plot such a parameter against the Martinelli parameter, X, which is a function of the ratio of the pressure drop over an incremental length of the conduit attributable to the liquid phase to the pressure drop over an incremental length of the conduit attributable to the vapor phase. Although the relationship between the quality of the fluid and the Martinelli parameter, X, is complex, high values of quality (mostly vapor) will correspond to low values of the Martinelli parameter and vice-versa.
One useful parameter for understanding the invention and the flow regimes depicted in FIG. 7 is the Froude number, Fr, which is a function of the mass velocity, quality, hydraulic diameter of the conduit, vapor and liquid densities and the acceleration, such as gravity field. The Froude number is well-understood and accepted in the art as an indicator, inter alia, of the ratio of shear forces to acceleration forces and will be used herein to include approximations of the Froude number, often referred to as a modified Froude number. The Froude number will vary directly with the quality and mass velocity and inversely with the square root of the gravitational field, vapor and liquid density and the hydraulic diameter of the conduit. If the Froude number, Fr, is graphed against the independent Martinelli parameter, X, the flow regime in any given portion of a condenser passage can be predicted. The various dotted and dashed lines of FIG. 7 depict the approximations of flow regime transition points published by different researchers, identified in the key contained in FIG. 7, and will be familiar to those skilled in the art.
The more important transitions shown in FIG. 7 are the curves attributed to Taitel and Dukler and Sardesai et al, both of which depict the approximate transition between stratified or gravity-dominated flow and annular or shear force dominated flow. These curves also divide the slug flow regime and approximate the transition between low-velocity or gravity dominated slug flow from high-velocity or shear force dominated slug flow. As indicated above in the description of different types of flow regimes, these curves and preferably the curve attributed to Sardesai et al are also considered to be reliable indicators of a sharp change in heat transfer efficiency.
It has also been long-recognized that the condensation of a vapor and, to a lesser extent, the reduction in temperature of a vapor reduces the volume of the vapor and, hence, the mass velocity. Therefore the Froude number may change with location within the conduit due to change in quality during condensation or change of mass velocity within the conduit or both. Since higher values of quality will be associated with lower values of Martinelli parameter, the flow from inlet to outlet within a condenser conduit will fall on some locus of points extending generally left to right across the flow regime map of FIG. 7. In a condenser application, Froude number would tend to decrease with increasing Martinelli parameter at least because of the change in fluid quality due to condensation as the fluid progresses through the length of the conduit.
Change of flow regime from one type of flow to another at any point in the passage may, in turn, cause deterioration or change of flow regime over the entirety of the passage. For instance, if slug flow were to occur at an end of the passage, due to decrease of vapor flow velocity and/or gravity, the resulting decrease in vapor flow rate would cause a change of flow regime throughout the passage. This is because deterioration of flow regime increases viscous drag of the fluid and a higher pressure would be required to maintain a given mass velocity once the flow regime has deteriorated at any point within the conduit. Such change of flow regime is unavoidable in condensers of high efficiency heat transfer systems since the greatest amount of heat will be absorbed during evaporation if the fluid is mostly or entirely liquid at the outlet of the condenser. Accordingly, the flow regime at the outlet of the condenser must be well within or even substantially toward the right side of the slug or plug flow areas depicted in FIG. 7 and the locus of points indicating the flow regime within the conduit will necessarily traverse the slug or plug flow areas of the flow regime map.
The use of tapered passages to alter flow rate is known, such as in U.S. Pat. Nos. 4,586,565 and 4,762,171, to Hallstrom et al. to compensate for changes in fluid volume and to adjust flow rate but such implementations have been directed to controlling flow velocity in evaporators. Velocity control has thus been accomplished by providing increasing area of the passages at the expense of reducing flow area in adjacent channels. These arrangements are therefore of greater than optimal volume and difficult to optimize since heat exchange rate will vary over the length of the passages due to the change of flow area in adjacent channels. It is also to be noted that the devices of the Hallstrom patents have a preferred operational orientation, implying sensitivity to the gravitational field and, hence, to acceleration forces of all kinds, including vibration.
The susceptibility to alteration of condenser performance by flow regime is especially critical in aircraft and aerospace applications where the direction and force of accelerations vary widely and gravity cannot be reliably exploited to enhance refrigerant flow. Moreover, in high performance aircraft, heat exchange requirements may be greatest at the same time when gravity or acceleration loads and power requirements other than for the compressor of the heat exchange system are high. Such applications also require light weight construction and compactness of the condenser structure.
Therefore a need exists for a condenser structure which will retain an efficient two phase flow regime and remain functionally unaffected under severe acceleration loads while being light, rugged and compact in construction and capable of being economically manufactured in a plurality of configurations for specific applications.